System and Method for Vibration in a Drill String

ABSTRACT

A system for damping vibration in a drill string can include a valve assembly having a supply of a fluid, a first member, and a second member capable of moving in relation to first member in response to vibration of the drill bit. The first and second members define a first and a second chamber for holding the fluid. Fluid can flow between the first and second chambers in response to the movement of the second member in relation to the first member. The valve assembly can also include a coil or a valve for varying a resistance of the fluid to flow between the first and second chambers.

CROSS REFERENCE TO RELATED APPLICATION

This application claims priority under 35 U.S.C.§119(e) to U.S.provisional application No. 60/518,116, filed Nov. 7, 2003, the contentsof which is incorporated by reference herein in its entirety.

Pursuant to 35 U.S.C.§202(c), it is acknowledged that the U.S.government may have certain rights to the invention described herein,which was made in part with funds from the Deep Trek program of the U.S.Department of Energy National Energy Technology Laboratory, Grant NumberDE-FC26-02NT41664.

FIELD OF THE INVENTION

The present invention relates to underground drilling, and morespecifically to a system and a method for damping vibration that occursin a drill string during drilling operations.

BACKGROUND OF THE INVENTION

Underground drilling, such as gas, oil, or geothermal drilling,generally involves drilling a bore through a formation deep in theearth. Such bores are formed by connecting a drill bit to long sectionsof pipe, referred to as a “drill pipe,” so as to form an assemblycommonly referred to as a “drill string.” The drill string extends fromthe surface to the bottom of the bore.

The drill bit is rotated so that the drill bit advances into the earth,thereby forming the bore. In rotary drilling, the drill bit is rotatedby rotating the drill string at the surface. Piston-operated pumps onthe surface pump high-pressure fluid, referred to as “drilling mud,”through an internal passage in the drill string and out through thedrill bit. The drilling mud lubricates the drill bit, and flushescuttings from the path of the drill bit. The drilling mud then flows tothe surface through an annular passage formed between the drill stringand the surface of the bore.

The drilling environment, and especially hard rock drilling, can inducesubstantial vibration and shock into the drill string. Vibration alsocan be introduced by factors such as rotation of the drill bit, themotors used to rotate the drill string, pumping drilling mud, imbalancein the drill string, etc. Such vibration can result in premature failureof the various components of the drill string. Substantial vibrationalso can reduce the rate of penetration of the drill bit into thedrilling surface, and in extreme cases can cause a loss of contactbetween the drill bit and the drilling surface.

Operators usually attempt to control drill string vibration by varyingone or both of the following: the rotational speed of the drill string,and the down-hole force on the drill bit (commonly referred to as“weight-on-bit”). These actions often do not effectively reduce thevibrations. Reducing the weight-on-bit or the rotary speed of the drillbit usually reduces drilling efficiency. In particular, drill bitstypically are designed for a predetermined rotary speed range andweight-on-bit. Operating the drill bit off of its design point canreduce the performance and the service life of the drill bit.

So-called “shock subs” are sometimes used to dampen drill stringvibrations. Shock subs, however, typically are optimized for oneparticular set of drilling conditions. Operating the shock sub outsideof these conditions can render the shock sub ineffective, and in somecases can actually increase drill string vibrations. Moreover, shocksubs and isolators usually isolate the portions of the drill stringup-hole of the shock sub or isolator from vibration, but can increasevibration in the down-hole portion of the drill string, including thedrill bit.

An ongoing need therefore exists for a system and method that can dampendrill-string vibrations, and particularly vibration of the drill bit,throughout a range of operating conditions.

SUMMARY OF THE INVENTION

A preferred embodiment of a valve assembly for damping vibration of adrill bit comprises a first member capable of being mechanically coupledto the drill bit so that the first member is subjected to vibration fromthe drill bit, and a supply of magnetorheological fluid.

The valve assembly also comprises a second member mechanically coupledto the first member so that the second member can translate in relationto the first member along a longitudinal centerline of the valveassembly, the first and second members defining a first and a secondchamber for holding the magnetorheological fluid. The first and secondchambers are in fluid communication.

The valve assembly further comprises a coil proximate to one the firstand the second members so that the magnetorheological fluid can besubjected to a magnetic field generated by the coil.

A preferred embodiment of a valve assembly for damping vibration of adrill bit in a drill string comprises a supply of a fluid, a firstmember capable of being coupled to the drill string so that the firstmember is subjected to vibration from the drill bit, and a second membercapable of moving in relation to first member in response to thevibration of the drill bit.

The first and second members define a first and a second chamber forholding the fluid. The first and second chambers are in fluidcommunication so that the fluid flows between the first and secondchambers in response to the movement of the second member in relation tothe first member. The valve assembly also comprises means for varying aresistance of the fluid to flow between the first and second chambers.

A preferred embodiment of a torsional bearing assembly for transmittingtorque to a drill bit comprises a first member capable of beingmechanically coupled to a source of the torque so that the first memberrotates in response to the torque. The first member has a first grooveformed therein.

The torsional bearing assembly also comprises a second member capable ofbeing mechanically coupled to the drill bit so that the drill bitrotates in response to rotation of the second member. The second memberis mechanically coupled to the first member so that the second membercan translate in relation to the first member in a first directionsubstantially coincident with a longitudinal centerline of the torsionalbearing assembly. The second member has a second groove formed thereinthat faces the first groove so that the first and second grooves form apassage extending substantially in a second direction.

The torsional bearing assembly also comprises a ball bearing disposed inthe passage for transmitting the torque between the first and the secondmembers.

A preferred embodiment of a spring assembly for use in a drill stringcomprises a first member capable of being mechanically coupled to thedrill bit so that the first member can translate in a first and anopposing second direction in response to the movement of the drill bit.

The spring assembly also comprises a second member mechanically coupledto the first member so that the first member can translate in relationto the second member in the first and the second directions, and aspring stack disposed on one of the first and the second members.

A first end of the spring stack is substantially restrained and a secondend of the spring translates in the first direction when the firstmember translates in the first direction in relation to the secondmember so that the spring stack is compressed. A second end of thespring stack is substantially restrained and the first end of the springstack translates in the second direction when the first membertranslates in the second direction in relation to the second member sothat the spring stack is compressed.

A preferred embodiment of a vibration damping system for use in a drillstring for drilling a drill hole comprises a bearing comprising a firstmember and a second member coupled to the first member so that the firstmember can translate in an up-hole and a down-hole direction in relationto the second member and torque can be transferred between the first andthe second members.

The vibration damping system also comprises a valve assembly comprisinga first member securely coupled to the first member of the torsionalbearing assembly so that the first member of the valve assemblytranslates in the up-hole and down-hole directions with the first memberof the torsional bearing assembly.

The valve assembly also comprises a second member securely coupled tothe second member of the torsional bearing assembly so that the secondmember of the valve assembly translates in the up-hole and down-holedirections with the second member of the first torsional bearingassembly, the first and second members of the valve assembly defining afirst and a second chamber for holding a supply of a fluid so that thefluid flows between the first and the second chambers in response torelative movement between the first and second members of the valveassembly. The valve assembly further comprises means for varying a flowresistance of the fluid.

The vibration damping system further comprises a spring assemblycomprising a first member securely coupled to the first member of thevalve assembly so that the first member of the spring assemblytranslates in the up-hole and down-hole directions with the first memberof the valve assembly.

The spring assembly also comprises a second member securely coupled tothe second member of the valve assembly the so that the second member ofthe spring assembly translates in the up-hole and down-hole directionswith the second member of the valve assembly. The spring assemblyfurther comprises a spring for resisting relative movement between thefirst and second members of the spring assembly.

A preferred method for damping vibration of a drill bit comprisesproviding a valve assembly capable of exerting a viscous damping forceon the drill bit, and controlling the viscous damping force in responseto at least one operating parameter of the drill bit.

BRIEF DESCRIPTION OF THE DRAWINGS

The foregoing summary, as well as the following detailed description ofa preferred embodiment, are better understood when read in conjunctionwith the appended diagrammatic drawings. For the purpose of illustratingthe invention, the drawings show an embodiment that is presentlypreferred. The invention is not limited, however, to the specificinstrumentalities disclosed in the drawings. In the drawings:

FIG. 1 is a longitudinal cross-sectional view of a preferred embodimentof a vibration damping system installed as part of a drill string;

FIG. 2 is a longitudinal cross-sectional view of a turbine alternatorassembly of the drill string shown in FIG. 1;

FIG. 3 is a longitudinal cross-sectional view of a torsional bearingassembly of the vibration damping system shown in FIG. 1;

FIG. 4 is a magnified view of the area designated “A” in FIG. 1;

FIG. 5 is a side view of a mandrel of the torsional bearing assemblyshown in FIG. 3;

FIG. 6 is a cross-sectional view taken through the line “B-B” of FIG. 3;

FIG. 7 is a magnified view of the area designated “C” in FIG. 3;

FIG. 8 is a magnified view of the area designated “D” in FIG. 3;

FIG. 9 is a side view of a reciprocating seal of the torsional bearingassembly shown in FIG. 3;

FIG. 10 is a longitudinal cross-sectional view of a valve assembly ofthe vibration damping system shown in FIG. 1;

FIG. 11 is a perspective view of a mandrel of the valve assembly shownin FIG. 10;

FIG. 12 is a block diagram of a controller for the valve assembly shownin FIG. 10;

FIG. 13 is a flow diagram for depicting a process for controlling anamount of damping developed by the valve assembly shown in FIG. 10;

FIG. 14 is a longitudinal cross-sectional view of a spring assembly ofthe vibration damping system shown in FIG. 1;

FIGS. 15A and 15B list equations for calculating a combined springconstant of a first and a second spring of the spring assembly shown inFIG. 14;

FIG. 16A is a longitudinal cross-sectional view of an alternativeembodiment of the valve assembly shown in FIG. 10, depicting a mandrelof the valve assembly in a neutral position;

FIG. 16B is a longitudinal cross-sectional view of the valve assemblyshown in FIG. 16A, depicting the mandrel in a position removed from itsneutral position;

FIG. 17A is a longitudinal cross-sectional view of another alternativeembodiment of the valve assembly shown in FIG. 10, depicting a mandrelof the valve assembly in a neutral position;

FIG. 17B is a longitudinal cross-sectional view of the valve assemblyshown in FIG. 17A, depicting the mandrel in a position removed from itsneutral position;

FIG. 18 is a longitudinal cross-sectional view of another alternativeembodiment of the valve assembly shown in FIG. 10;

FIG. 19 is a cross-sectional side view of the valve assembly shown inFIG. 10, showing lines of magnetic flux generated by coils of the valveassembly; and

FIG. 20 depicts a curve of desired damping as a function ofdisplacement, for the valve assembly shown in FIGS. 10 and 19.

DESCRIPTION OF PREFERRED EMBODIMENTS

The figures depict a preferred embodiment of a vibration damping system10. The figures are each referenced to a common coordinate system 11depicted therein. The vibration damping system 10 can be used as part ofa drill string 12, to dampen vibration of a drill bit 13 located at adown-hole end of the drill string 12 (see FIG. 1).

The vibration damping system 10 comprises a torsional bearing assembly14, a valve assembly 16, and a spring assembly 18. The valve assembly 16and the spring assembly 18 can produce axial forces that dampenvibration of the drill bit 13. The magnitude of the damping force can bevaried by the valve assembly 14 in response to the magnitude andfrequency of the vibration, on a substantially instantaneous basis. Thevibration damping assembly 10 can be mechanically coupled to the drillbit by drill pipe 22 that forms part of the drill string 12.

The torsional bearing assembly 14 can facilitate the transmission ofdrilling torque, while permitting relative axial movement between theportions of the drill string 12 located up-hole and down-hole of thevibration damping system 10. Moreover, the torsional bearing assembly 14can transform torsional vibration of the drill bit 13 into axialvibration. The axial vibration, in turn, can be damped by the valveassembly 16 and the spring assembly 18.

The vibration damping system 10 can be mechanically and electricallyconnected to a turbine-alternator module 20 located up-hole of thevibration damping system 10 (see FIGS. 1 and 2). (The up-hole anddown-hole directions correspond respectively to the “+x” and “−x”directions denoted in the figures.) The turbine-alternator module 20 canprovide electric power for the vibration damping system 10. The use ofthe vibration damping system 10 in conjunction with theturbine-alternator module 20 is described for exemplary purposes only.The vibration damping system 10 can be powered by an alternative meanssuch as a battery located in the vibration damping system 10 (orelsewhere in the drill string 12), or a power source located aboveground.

The torsional bearing assembly 14 comprises a casing 50 and a bearingmandrel 52 (see FIGS. 3, 5, and 6). The bearing casing 50 and thebearing mandrel 52 are disposed in a substantially coaxial arrangement,with the bearing mandrel 52 located within the bearing casing 50. Thebearing mandrel 52 is supported within the bearing casing 50 by a radialbearing 54. The bearing casing 50 can translate axially in relation tothe bearing mandrel 52. The torsional bearing assembly 12 also comprisesa plurality of ball bearings 55 for transmitting torque between thebearing mandrel 52 and the bearing casing 50. The ball bearings 55 canbe, for example, rock bit balls (other types of ball bearings can beused in the alternative).

Drilling torque is transmitted to an outer casing 21 of theturbine-alternator module 20 by way of a drill pipe 22 located up-holeof the turbine-alternator module 20 (see FIG. 1). The bearing mandrel 52is secured to the outer casing 21 so that the drilling torque istransferred to the bearing mandrel 52. The bearing mandrel 52 thereforerotates, and translates axially with the outer casing 21.

A centralizer feed-thru 56 is positioned within the bearing mandrel 52,proximate the up-hole end thereof, and is secured to the bearing mandrel52 by a locking pin 57 (see FIGS. 1 and 4). The centralizer feed-thru 56can be supported by one or more ribs (not shown).

The centralizer feed-thru 56 facilitates routing of electrical signalsand power between the turbine-alternator assembly 20 and the torsionalbearing assembly 12. In particular, the centralizer feed-thru 56includes a multi-pin connector 58 for electrically connecting thecentralizer feed-thru 56 to the turbine-alternator assembly 20. Thecentralizer feed-thru 56 also includes a second electrical connector 59.Wiring (not shown) is routed from the connector 58 to the connector 59by way of a passage 60 formed within the centralizer feed-thru 56.(Additional wiring (also not shown) is routed from the electricalconnector 59 and through a wireway formed in the bearing mandrel 52.)The centralizer feed-thru 56 also includes a removable panel 60 forproviding access to the locking pin 57 and the connector 59.

The centralizer feed-thru 56 has a passage 61 formed therein. Thepassage 61 adjoins a passage 63 defined in the bearing mandrel 52 by aninner surface 64 thereof. The passage 63 receives drilling mud from thepassage 61.

The bearing mandrel 52 has a plurality of grooves 70 formed in an outersurface 72 thereof (see FIG. 5). The grooves 70 are substantiallyparallel, and are spaced apart in substantially equal angular incrementsalong the outer surface 72. (The grooves 70 can be spaced apart inunequal angular increments in alternative embodiments.) The surfaces ofthe bearing mandrel 52 that define the grooves 70 each havesubstantially semi-circular shape, to accept the substantially sphericalball bearings 55.

The depth of each groove 70 is substantially constant along the lengththereof. Preferably, the grooves 70 are substantially straight. In otherwords, a longitudinal centerline 80 of each groove 70 is shapedsubstantially as a helix.

The bearing casing 50 has a plurality of grooves 74 formed on an innersurface 76 thereof (see FIGS. 3, 5, and 6). The size, shape, andorientation of the grooves 74 are approximately equal those of thegrooves 70.

Each groove 74 faces a corresponding one of the grooves 70 when thebearing casing 50 and the bearing mandrel 52 are assembled. Eachcorresponding groove 70 and groove 74 define a passage 78 for ten of theball bearings 55 (see FIG. 3). Each passage 78 preferably has a lengthgreater than a combined length of the ten ball bearings 55 disposedtherein, to facilitate translation of the ball bearings 50 along thepassage 78. (The number of ball bearings 55 within each groove 70 isapplication dependent, and can vary with factors such as the amount oftorque to be transferred between the bearing casing 50 and the bearingmandrel 52; more or less than ten of the ball bearings 55 can bedisposed in each groove 70 in alternative embodiments.)

The grooves 70 and the grooves 74 are sized so that sufficient clearanceexists between the walls of the grooves 70, 74 and the associated ballbearings 55 to permit the ball bearings 55 to translate in thelengthwise direction within the passages 78.

Each groove 70 preferably is angled in relation to a longitudinalcenterline 82 of the bearing mandrel 52 (see FIG. 5). (Axially-alignedgrooves can be used in the alternative, for reasons discussed below.)(The longitudinal centerline 82 of the bearing mandrel 52 is orientedsubstantially in the axial (“x”) direction). In particular, a centerline80 of each groove 70 is oriented in relation to the centerline 82 at ahelix angle denoted by the reference symbol “∃” in FIG. 5. Preferably,the helix angle ∃ lies within a range of approximately four degrees toapproximately fifteen degrees.

The optimal value for the helix angle ∃ is application dependent; aparticular value is presented for exemplary purposes only. Inparticular, the optimal value for ∃ can be calculated based on thefollowing parameters: maximum torque (T) and maximum allowable axialforce (F_(A)) to be transmitted through the drill string 12; radialdistance (R) between the centerline 82 of the bearing mandrel 52 and thecenters of the ball bearings 55; and maximum tangential force (F_(C)) onthe ball bearings 55 (equal to T/R). The helix angle∃=arcsine(F_(A)/F_(C)).

Drilling torque transmitted to the bearing mandrel 52 from theturbine-alternator assembly 20 exerts a tangential force, i.e., a forcecoincident with the “y-z” plane, on the ball bearings 55. The tangentialforce is transferred to the ball bearings 55 by way of the walls of thegrooves 70. The ball bearings 55 transfer the torque to the bearingcasing 50 by way of the walls of the grooves 74, thereby causing thebearing casing 50 to rotate with the bearing mandrel 52.

Movement of the ball bearings 55 along the length of their respectivepassage 80 can facilitate relative movement between the bearing mandrel52 and the bearing casing 50 in the axial direction. Hence, thetorsional bearing assembly 14 substantially decouples the portion of thedrill string 12 down-hole of the vibration damping system 10 from axialmovement of the portion of the drill string 12 up-hole of the vibrationdamping system 10, and vice versa.

The use of the ball bearings 55 is believed to minimize friction, andthe sticking associated therewith, as the bearing mandrel 52 translatesaxially in relation to the bearing casing 50. Alternative embodimentscan be configured with other means for facilitating relative axialmovement between the bearing mandrel 52 and the bearing casing 50.

The bearing mandrel 52 and the bearing casing 50 are restrained fromrelative tangential movement, i.e., movement in the “y-z” plane, due tothe substantially straight geometry of the passages 78, and because theball bearings 55 remain at a substantially constant distance from thecenterline 82 of the bearing mandrel 52 as the ball bearings 57translate along their associated passages 78.

The bearing casing 50 is connected to the drill bit 13 by way of thevalve assembly 16, the spring assembly 18, and the portion of the drillstring 12 located down-hole thereof. The bearing casing 50 thereforerotates with the drill bit 13, and translates with the drill bit 13 inthe axial direction. Hence, axial and torsional vibrations of the drillbit 13 are transmitted up-hole by way of the drill string 12, to thebearing casing 50.

Orienting the passages 78 at the helix angle ∃, it is believed, cantransform at least a portion of the torsional vibration acting on thebearing casing 50 into axial vibration. In particular, the angledorientation of the passages 78 permits the bearing casing 50 to rotate(by a minimal amount) in relation to the bearing mandrel 52 in responseto torsional vibration. The rotation of the bearing casing 50 isconverted into an axial force due to the angled orientation of thepassages 78. Hence, the torsional vibration acting on the bearing casing50 can be converted, at least in part, into axial vibration acting onthe bearing mandrel 52. This axial vibration, as discussed below, can betransferred to and damped by the valve assembly 16 and the springassembly 18. (In addition, the angled orientation of the passages 78 isbelieved to generate friction damping that further reduces the torsionalvibration.)

It should be noted that the grooves 70, 74 in alternative embodimentscan be formed so that the passages 70 extend in a directionsubstantially parallel to the longitudinal centerline 82 of the bearingmandrel 52. (Torsional vibration of the drill bit 13 will not beconverted into axial vibration in the above-described manner, in thesetypes of embodiments.)

The torsional bearing assembly 14 also comprises a linear variabledisplacement transducer (LVDT) 84 for measuring the relativedisplacement of the bearing casing 50 and the bearing mandrel 52 in theaxial direction (see FIGS. 3 and 7). The LVDT 84 comprises an array ofaxially-spaced magnetic elements 86 embedded in the bearing casing 50,proximate the inner surface 76 thereof. The LVDT 84 also comprises asensor 88, such as a Hall-effect sensor, mounted on the bearing mandrel52 so that the sensor 88 is magnetically coupled to the magneticelements 86.

The sensor 88 produces an electrical output as a function of theposition of the sensor 88 in relation to the array of magnetic elements86. The LVDT 84 thereby can provide an indication of the relative axialpositions of the bearing casing 50 and the bearing mandrel 52. Moreover,the rate of change of the output is a function of the rate of change inthe relative positions of the sensor 88 and the array of magneticelements 86. Hence, the LVDT 84 can provide an indication of therelative axial displacement, velocity, and acceleration of the bearingcasing 50 and the bearing mandrel 52.

The torsional bearing assembly 14 also includes a compensation piston 90(see FIGS. 3 and 8). As shown best in FIG. 8, the compensation piston 90is positioned between the bearing mandrel 52 and the bearing casing 50,proximate an up-hole end of the bearing casing 50. An up-hole side 90′of the compensation piston 90 is exposed to drilling mud. A down-holeside 90″ of the compensation piston 90 is exposed to compensation oilused to equalize the pressurize within the interior of the vibrationdamping system 10.

The compensation piston 90 can slide in the axial direction in relationto the bearing casing 50 and the bearing mandrel 52, in response to apressure differential between the drilling mud and the compensation oil.This feature can to help to equalize the pressure between thecompensation oil and the drilling mud, and compensate for thermalexpansion of the compensation oil. In particular, the movement of thecompensation piston 90 can help to pressurize the compensation oil asthe distance of the drill bit 13 below ground level increases (therebycausing an increase in the pressure of the drilling mud).

Three reciprocating seals 91 are positioned in grooves 92 formed aroundthe outer circumference of the compensation piston 90 (see FIGS. 3, 8,and 9). The seals 91 substantially isolate the compensation oil from thedrilling mud. Two of the seals 91 preferably face the drilling mud, soas to discourage infiltration of the drilling mud into the compensationoil.

Each seal 91 includes a heel 93, a lip (scraper) 94, and an extension95. The lip 94 adjoins the heel 93, and forms part of the inner diameterof the seal 91. The extension 95 adjoins the heel 93, and forms part ofthe outer diameter of the seal 91. The heel 93, lip 94, and extension 95preferably are formed from a wear and extrusion-resistant material, suchas a blend of polytetrafluoroethylene (PTFE) and carbon-graphite.

The heel 93, lip 94, and extension 95 define a groove 96. A spring 97 isdisposed in the groove 96. The spring 97 preferably is a ribbon spring.Preferably, the spring 97 is formed from a resilient,corrosion-resistant material such as Elgiloy. The spring 97 exerts aforce on the lip 94 in the radially-outward direction. The force urgesthe lip 94 into contact with the adjacent surface of the bearing mandrel52, and can help to maintain this contact as the lip 94 wears.

The groove 96 preferably is sized so that the surface area of the seal91 that defines the groove 96 is minimal This feature can help tominimize the pressure forces exerted on the lip 94 by the drilling mudor the compensation oil.

The geometry of the lip 94, it is believed, causes the lip 94 to scrape(rather than slide over) the drilling mud or the compensation oil on theadjacent surface of the bearing mandrel 52 as the compensation piston 90translates in relation thereto (the seals 91 therefore are believed tobe particularly well suited for use with an abrasive materials such asdrilling mud or magnetorheological fluid).

The extension 95 helps to maintain spacing between the lip 94, and thegap between the bearing mandrel 52 and the compensation piston 90. Thisfeature therefore can reduce the potential for the lip 94 to becometrapped in the gap and damaged during movement of the compensationpiston 90.

The heel 93 preferably is sized so that the height of the seal 91exceeds the height of the corresponding groove 92. The seals 91therefore can act as glide rings that support the compensation piston 90on the bearing mandrel 52.

The relatively large size of the heel 93 is believed to help the heel 93resist the potentially large differential pressures that can form acrossthe seal 91.

The valve assembly 16 is located immediately down-hole of the torsionalbearing assembly 12 (see FIGS. 1 and 10). The valve assembly 16comprises a valve casing 102. The valve casing 102 comprises an outercasing 103, and a housing 104 positioned within the outer casing 103.

The valve assembly 16 also comprises a coil mandrel 106 positionedwithin the valve casing 102 (see FIGS. 10 and 11). The outer casing 103,housing 104, and coil mandrel 106 are disposed in a substantiallycoaxial arrangement. The coil mandrel 106 preferably is formed from amaterial having a high magnetic permeability and a low magneticsusceptibility, such as 410 stainless steel.

The coil mandrel 106 is secured to the bearing mandrel 52 so that thecoil mandrel 106 rotates, and translates axially with the bearingmandrel 52.

As shown in FIG. 3, the outer portion 103 of the valve casing 102 issecured to the bearing casing 50 so that the drilling torque istransferred from the bearing casing 50 to the valve casing 102. Thevalve casing 102 therefore rotates, and translates axially with thebearing casing 50.

The housing 104 preferably comprises a first portion 108, and a secondportion 110 located down-hole of the first portion 108 (see FIG. 10).The housing 104 also comprises a third portion 112 located down-hole ofthe second portion 110. (It should be noted that the housing 104 can beformed as one piece in alternative embodiments. Moreover, the housing104 and the outer casing 103 can be formed as one piece in alternativeembodiments.)

The up-hole end of the first portion 108 abuts a lip (not shown) on theouter casing 103 of the valve casing 102. The down-hole end of the thirdportion 112 abuts a radial bearing 120 of the valve assembly 16 (seeFIG. 10). This arrangement restrains the housing 104 from axial (“x”direction) movement in relation to the outer casing 103. (The housing104 therefore translates axially with the outer casing 103.)

The valve assembly 16 also comprises a sleeve 122 (see FIG. 10). Thesleeve 122 is concentrically disposed around portion of the coil mandrel106, proximate the down-hole end thereof. The sleeve 122 is secured tothe coil mandrel 106 so that the sleeve 122 rotates, and translatesaxially with the coil mandrel 106.

A first linear bearing 125 is positioned in a groove formed around thecoil mandrel 106, proximate the up-hole end thereof. A second linearbearing 126 is positioned in a groove formed around the sleeve 122. Thefirst and second linear bearings 125, 126 help to support the coilmandrel 106 and the sleeve 122, and facilitate axial movement of thecoil mandrel 106 and the sleeve 122 in relation to the housing 104 (andthe valve casing 102).

An inner surface 124 of the coil mandrel 106 defines a passage 127 forpermitting drilling mud to flow through the valve assembly 16. Thepassage 127 adjoins the passage 63 formed in the bearing mandrel 52.

The coil mandrel 106 has a plurality of outwardly-facing recesses 128formed around a circumference thereof (see FIGS. 10 and 11). Adjacentones of the recesses 128 are separated by outer surface portions 130 ofthe coil mandrel 106.

The coil mandrel 106 and the second portion 110 of the housing 104 aresized so that a clearance, or gap 135 exists between an inner surface132 of the second portion 110, and the adjacent outer surface portions130 of the coil mandrel 106 (see FIG. 10). The gap 135 preferably iswithin the range of approximately 0.030 inch to approximately 0.125inch. (The optimal value, or range of values for the gap 135 isapplication-dependent; a specific range of values is presented forexemplary purposes only.)

The valve assembly 16 also comprises a plurality of coils 136. Each ofthe coils 136 is wound within a respective one of the recesses 128.Adjacent ones of the coils 136 preferably are wound in oppositedirections (the purpose of this feature is discussed below).

A groove 140 is formed in each of the outer surface portions 130 tofacilitate routing of the wiring for the coils 136 between adjacent onesof the recesses 128 (see FIG. 11). The grooves 140 each extendsubstantially in the axial (“x”) direction. A wireway 142 and anelectrical feed thru 144 are formed in the coil mandrel 106 tofacilitate routing of the wire 138 from the up-hole end of the coilmandrel 106 to the recesses 128 (see FIG. 10). (The coils 136 can bepositioned on the valve casing 102 instead of (or in addition to) thecoil mandrel 106 in alternative embodiments.)

The coils 136 each generate a magnetic field 149 in response to thepassage of electrical current therethrough (the magnetic fields 149 aredepicted diagrammatically in FIG. 19). The coils 136 can be electricallyconnected to a controller 146 mounted in the turbine-alternator assembly20 (see FIG. 2). The controller 146 can be powered by an alternator 147of the turbine-alternator assembly 20. The controller 146 can supply anelectrical current to the coils 136. The controller 146 can control themagnitude of the electrical current to vary the strength of theaggregate magnetic field generated by the coils 136. Further detailsrelating to this feature are presented below.

The controller 146 is depicted as being mounted within theturbine-alternator assembly 20 for exemplary purposes only. Thecontroller 146 can be mounted in other locations, including above-groundlocations, in the alternative.

The first portion 108 of the housing 104 and the coil mandrel 106 definea circumferentially-extending first, or up-hole, chamber 150 (see FIG.10). The third portion 112 of the housing 104 and the coil mandrel 106define a circumferentially-extending second, or down-hole chamber 152.

The first and second chambers 150, 152 are filled with amagnetorheological fluid (hereinafter referred to as “MRF”). MRFstypically comprise non-colloidal suspensions of ferromagnetic orparamagnetic particles. The particles typically have a diameter greaterthan approximately 0.1 microns. The particles are suspended in a carrierfluid, such as mineral oil, water, or silicon.

Under normal conditions, MRFs have the flow characteristics of aconventional oil. In the presence of a magnetic field (such as themagnetic fields 149), however, the particles suspended in the carrierfluid become polarized. This polarization cause the particles to becomeorganized in chains within the carrier fluid.

The particle chains increase the fluid shear strength (and therefore,the flow resistance or viscosity) of the MRF. Upon removal of themagnetic field, the particles return to an unorganized state, and thefluid shear strength and flow resistance returns to its previous value.Thus, the controlled application of a magnetic field allows the fluidshear strength and flow resistance of an MRF to be altered very rapidly.MRFs are described in U.S. Pat. No. 5,382,373 (Carlson et al.), which isincorporated by reference herein in its entirety. An MRF suitable foruse in the valve assembly 16 is available from APS Technology ofCromwell, Conn.

The first chamber 150 and the second chamber 152 are in fluidcommunication by way of the gap 135 formed between the inner surface 132of the second portion 110, and the adjacent outer surface portions 130of the coil mandrel 106. Hence, the MRF can move between the first andsecond chambers 150, 152 by way of the gap 135.

The MRF in the first chamber 150 is substantially isolated from thecompensation oil located up-hole thereof by three of the reciprocatingseals 91 (as described above in relation to the compensation piston 90)disposed in grooves formed in the coil mandrel 106. The MRF in thesecond chamber 152 is substantially isolated from the compensation oillocated down-hole thereof by three more of the seals 91 disposed inadditional grooves formed in the sleeve 122. Two of the seals 91 in eachset of three face the MRF in the associated chamber 150, 152, todiscourage infiltration of the MRF into the chamber 150, 152.

The outer portion 103 of the valve casing 102 is connected to the drillbit 13 by way of the spring assembly 18 and the portion of the drillpipe 22 located down hole of the vibration damping system 10. The outerportion 103 therefore rotates, and translates axially with the drill bit13. Moreover, the coil mandrel 106 and the sleeve 122 are substantiallydecoupled from axial movement of the valve casing 102 by the torsionalbearing assembly 14.

The above-noted arrangement causes the coil mandrel 106 and the sleeve122 to reciprocate within the housing 104 in response to vibration ofthe drill bit 13. This movement alternately decreases and increases therespective volumes of the first and second chambers 150, 152. Inparticular, movement of the coil mandrel 106 and the sleeve 122 in theup-hole direction in relation of the housing 104 increases the volume ofthe first chamber 152, and decreases the volume of the second chamber150. Conversely, movement of the coil mandrel 106 and the sleeve 122 inthe down-hole direction in relation of the housing 104 decreases thevolume of the first chamber 152, and increases the volume of the secondchamber 150. The reciprocating movement of the coil mandrel 106 and thesleeve 122 within the housing 104 thus tends to pump the MRF between thefirst and second chambers 150, 152 by way of the gap 135.

The flow resistance of the MRF causes the valve assembly 16 to act as aviscous damper. In particular, the flow resistance of the MRF causes theMRF to generate a force (opposite the direction of the displacement ofthe coil mandrel 106 and the sleeve 122 in relation to the housing 104)that opposes the flow of the MRF between the first and second chambers150, 152. The MRF thereby resists the reciprocating motion of the coilmandrel 106 and the sleeve 122 in relation to the housing 104. Thisresistance can dampen axial vibration of the drill bit 13.

The magnitude of the damping force generated by the MRF is proportionala function of the flow resistance of the MRF and the frequency of theaxial vibration. The flow resistance of the MRF, as noted above, can beincreased by subjecting the MRF to a magnetic field. Moreover, the flowresistance can be varied on a substantially instantaneous basis byvarying the magnitude of the magnetic field.

The coils 136 are positioned so that the lines of magnetic fluxgenerated by the coils 136 cut through the MRF located in the first andsecond chambers 150, 152 and the gap 135 (see FIG. 19). The currentthrough the coils 136, and thus the magnitude of the magnetic flux, iscontrolled by the controller 146. The use of multiple axially-spacedcoils 136 is believed to distribute the magnetic fields 149 axiallywithin the MRF, helping to ensure that the MRF is exposed to themagnetic flux regardless of the position of the coil mandrel 106 inrelation to the housing 104 and the valve casing 102. Distributing themagnetic fields 149 in this manner thus can help to maximize the dampingforce by energizing a greater percentage of the MRF.

The controller 146 can control the current (power) through the coils 136in response to vibration of the drill bit 13 so as to dampen vibrationof the drill bit 13 (the process by which the controller performs thisfunction is depicted in the form of a flow diagram in FIG. 130.

The controller 146 preferably comprises a computing device 160 (see FIG.12.) The computing device 160 can be, for example, a programmablemicroprocessor such as a digital signal processing (DSP) chip. Thecontroller 146 also comprises a memory storage device 162, solid staterelays 162, and a set of computer-executable instructions 164. Thememory storage device 162 and the solid state relays 162 areelectrically coupled to the computing device 160, and thecomputer-executable instructions 164 are stored on the memory storagedevice 162.

The controller 146 is configured as a printed circuit board mounted inthe turbine-alternator module 20. The controller 146 can be configuredin other ways in alternative embodiments.

The LVDT 84 is electrically connected to the computing device 160. TheLVDT 84 provides an input to the computing device 160 in the form of anelectrical signal indicative of the relative axial position, velocity,and acceleration of the bearing casing 50 and the bearing mandrel 52, asnoted above. The bearing casing 50 is connected the drill bit 12, and issubstantially decoupled from axial movement of the bearing mandrel 52.Hence, the output of the LVDT 84 is responsive to the magnitude andfrequency of the axial vibration of the drill bit 13.

The computer executable instructions 164 include algorithms that candetermine the optimal amount of damping at a particular operatingcondition, based on the output of the LVDT 84, i.e., based on thedisplacement of the bearing mandrel 52 in relation to the bearing casing50.

It is believed that the optimal damping level increases with thedisplacement of the bearing mandrel 52 in relation to the bearing casing50. Moreover, lighter weight on bit conditions are believed to requireless damping than higher weight on bit conditions. Also, the optimalamount of damping is believed to increase with the stroke of the bearingmandrel 52 in relation to the bearing casing 50.

The desired damping at a particular condition can be calculated asfollows:

c=A×d ^(n) +B

where:

-   -   c=required damping (lb-sec/in)    -   d=relative displacement (as measured by the LVDT 84)    -   n =defines the shape of the damping curve    -   A =(damping_(max)−damping_(min))/disp^(n)    -   Damping_(max)=The maximum damping that occurs at the maximum        displacement    -   Damping_(min)=The minimum damping that occurs at the minimum        displacement or neutral point of the tool    -   Disp.=maximum relative displacement (4-inches, for example, for        the valve assembly 14)    -   B=min. damping

The desired damping of the valve assembly 14 is presented as a functionof displacement (as measured by the LVDT 84) in FIG. 20.

The desired damping also can be defined as a quadratic equation, or as alookup table in the controller 146.

The computer executable instructions 164 also determine the amount ofelectrical current that needs to be directed to the coils 136 to providethe desired damping. The controller 146 can process the input from theLVDT 84, and generate a responsive output in the form of an electricalcurrent directed to the coils 136 on a substantially instantaneousbasis. Hence, the valve assembly 16 can generate a damping force inresponse to vibration of the drill bit 13 on a substantiallyinstantaneous basis.

Preferably, the damping force prevents the drill bit 13 from losingcontact with the drilling surface due to axial vibration. The controller146 preferably causes the damping force to increase as the drill bit 13moves upward, to help maintain contact between the drill bit 13 and thedrilling surface. (Ideally, the damping force should be controlled sothe weight-on-bit remains substantially constant.) Moreover, it isbelieved that the damping is optimized when the dynamic spring rate ofthe vibration damping system 10 is approximately equal to the staticspring rate. (More damping is required when the dynamic spring rate isgreater than the static spring rate, and vice versa.)

It should be noted that alternative embodiments of the vibration dampingsystem 10 can include sensors in addition to, or in lieu of the LVDT 84.For example, the controller 146 can be programmed to determine therequisite damping based on inputs from one or more accelerometers,weight-on-bit sensors, velocity transducers, torque-on-bit sensors, etc.

The valve assembly 16 and the controller 146 can automatically increaseor decrease the amount of damping exerted on the drill bit 13 to reducevibration of the drill bit 13. The valve assembly 16 and the controller146 can perform this function on a substantially instantaneous basis, inresponse to one or more measured operating parameters. The ability toactively control vibration of the drill bit 13 in this manner, it isbelieved, can increase the rate of penetration of the drill bit, reduceseparation of the drill bit 13 from the drilling surface, lower orsubstantially eliminate shock on the drill bit, and increase the servicelife of the drill bit 13 and other components of the drill string 12.Moreover, the valve assembly 16 and the controller 146 can provideoptimal damping under variety of operating conditions, incontradistinction to shock subs. Also, the use of MRF to provide thedamping force makes the valve assembly 14 more compact than otherwisewould be possible.

The spring assembly 18 is located immediately down-hole of the valveassembly 16 (see FIGS. 1 and 14). The spring assembly 18 can exert arestoring force on the drill bit 13 in response to axial movement of thedrill bit 13 (the vibration damping assembly 10 thus behaves as aspring-mass-damper system).

The spring assembly 18 comprises a spring casing 200. The up-hole end ofthe spring casing 200 is secured to the outer casing 103 of the valvecasing 102 so that drilling torque is transferred to the spring casing200. The down-hole end of the spring casing 200 is secured to a casing302 of a compensation module 300, so that the drilling torque istransferred from the spring casing 200 to a casing 302 of thecompensation module 300. The spring casing 200 and the casing 302therefore rotate, and translate axially with the valve casing 102.

The spring assembly 18 also includes a spring mandrel 202, and a springstack 205. The spring stack 205 preferably comprises a first spring 206,and a second spring 208. (The spring stack 205 can include more or lessthan two springs in alternative embodiments.)

The spring casing 200, the spring mandrel 202, and the spring stack 205are disposed in a substantially coaxial relationship. The first andsecond springs 206, 208 are positioned in series, i.e., end to end,within the spring casing 202. The spring mandrel 202 is positionedwithin the first and the second springs 206, 208. (The relative axialpositions of the first and second springs 206, 208 can reversed fromthose depicted in FIG. 14, in alternative embodiments.)

The spring mandrel 202 can translate axially in relation to the springcasing 200. An inner surface 209 of the spring mandrel 202 defines apassage 210 for permitting drilling mud to flow through the springassembly 18.

The first and the second springs 206, 208 preferably are Bellevillesprings (other types of springs can be used in the alternative).Preferably, the second spring 208 is stiffer, i.e., has a higher springrate, than the first spring 206. This feature, as discussed below, isbelieved to facilitate transmission of axial vibration from the drillbit 13 to the valve assembly 14 under a relatively wide range ofweight-on-bit conditions. (Other spring configurations are possible inalternative embodiments. For example, one relatively soft Bellevillespring can be positioned between two relatively hard Belleville springsin one possible alternative embodiment.)

The compensation module 300 also includes a mandrel 304, and a slidingcompensation piston 306. The compensation piston 306 is positionedaround a down-hole portion of the mandrel 304.

The mandrel 304 of the compensation module 300 extends into thedown-hole portion of the spring casing 200. The mandrel 304 issupported, in part, by a radial bearing 305 positioned between themandrel 304 and the spring casing 200. A down-hole end of the bearing305 abuts an forward edge of the casing 302, thereby restraining thebearing 305 in the rearward direction.

An inner surface 310 of the mandrel 304 defines a passage 312 forpermitting drilling mud to flow through the mandrel 304 and into thecompensation module 300. The drilling mud, upon exiting the passage 312,enters a passage 314 defined by an inner surface 315 of the mandrel 304.(The drilling mud in the passage 314 acts against the down-hole side ofthe compensation piston 306.

The up-hole side of the compensation piston 306, the inner surface 310of the casing 302, and the mandrel 304 define acircumferentially-extending chamber 316 within the compensation module300. The chamber 316 is filed with compensating oil. Three of the seals91 are positioned in grooves formed in the compensation piston 306 sealthe chamber 316 to substantially isolate the compensation oil in thechamber 316 from the drilling mud in the passage 314. Two of the seals91 preferably face the drilling mud to discourage infiltration of thedrilling mud into the compensation oil.

The compensation piston 306 can slide in the axial direction in relationto the casing 302 and the mandrel 304, in response to a pressuredifferential between the compensation oil in the chamber 316, and thedrilling mud in the passage 314. This feature can to help to equalizethe pressure between the compensation oil and the drilling mud. Inparticular, the movement of the compensation piston 306 can help topressurize the compensation oil as the distance of the drill bit 13below ground level increases (thereby causing an increase in thepressure of the drilling mud).

It should be noted that details of the compensation module 300 arepresented for illustrative purposes only; the vibration damping system10 can be used in conjunction with other types of drill-stringcomponents located immediately down-hole thereof.

A coupling 211 is positioned within the spring casing 200, proximate anup-hole end thereof. The coupling 211 preferably has a substantiallyH-shaped cross section, as depicted in FIG. 14. The coupling 211receives the down-hole end of the coil mandrel 106, and the up-hole endof the spring mandrel 202. The coil mandrel 106 and the spring mandrel202 are secured to the coupling 211 so that the spring mandrel 202rotates, and translates axially with the coil mandrel 106.

A first spacer 212 is located immediately up-hole of the coupling 211,and separates the coupling 211 from the sleeve 122 of the valve assembly16.

A second spacer 214 is positioned between the coupling 211, and thefirst and spring 206. The first and the second springs 206, 208 urge thesecond spacer 214 into a lip 216 of the spring casing 200. Contactbetween the second spacer 214 and the lip 216 prevents movement of thesecond spacer 214 past the lip 216, and thereby restrains the first andsecond springs 206, 208 in the forward direction.

The rearward end of the spring mandrel 202 is positioned within themandrel 304 of the compensation module 300. The spring mandrel 202 andthe mandrel 304 can be secured by a suitable means such as aninterference fit. The mandrel 304 therefore rotates, and translateaxially with the spring mandrel 202.

A third spacer 218 is positioned between the second spring 208, themandrel 304, and the bearing 305. The first and the second springs 206,208 urge the third spacer 218 into the forward edge of the bearing 305.Contact between the third spacer 218 and the bearing 305 preventsmovement of third spacer 218 in the down-hole direction, and therebyrestrains the first and second springs 206, 208 in the down-holedirection.

The first and the second springs 206, 208 therefore are constrainedbetween the second and third spacers 214, 218. This arrangement causesthe first and second springs 206, 208 to function as double (dual)action springs. In particular, movement of the spring casing 200 in thedown-hole direction in relation of the spring mandrel 202 causes the lip216 of the spring casing 200 to urge the second spacer 214 in thedown-hole direction. (This type of relative movement can occur duringvibration-induced movement of the drill bit 13 in the down-holedirection.)

The second spacer 214, in turn, urges the first and second springs 206,208 in the down-hole direction, against the third spacer 218. The thirdspacer 218, in response, acts against the mandrel 304 of thecompensation assembly 300 in the down-hole direction. The mandrel 304,which is connected to the up-hole portion of the drill string 12 by wayof the spring mandrel 202, coil mandrel 106, and bearing mandrel 52,reacts the force exerted thereon by the third spacer 218.

The first and second springs 206, 208 therefore become compressed inresponse to the movement of the spring casing 200 in the down-holedirection. The resulting spring force acts against the spring casing 200(and the drill bit 13) in the up-hole direction, by way of the lip 216.The magnitude of the spring force is a function of the deflection of thespring casing 200 and the drill bit 13.

Movement of the spring casing 200 in the up-hole direction in relationof the spring mandrel 202 causes the forward edge of the casing 302(which is secured to the spring casing 200) to act against the bearing305. (This type of relative movement can occur during vibration-inducedmovement of the drill bit 13 in the up-hole direction.)

The bearing 305, in turn, urges the third spacer 218 and the adjacentfirst and second springs 206, 208 in the up-hole direction, toward thesecond spacer 214 and the coupling 211. The coupling 211, which isconnected to the up-hole portion of the drill string 12 by way of thespring coil mandrel 106 and the bearing mandrel 52, reacts the forceexerted thereon by the second spacer 214.

The first and second springs 206, 208 therefore become compressed inresponse to the movement of the spring casing 200 in the up-holedirection. The resulting spring force acts against the spring casing 200(and the drill bit 13) in the down-hole direction, by way of the bearing305 and the casing 302. The magnitude of the spring force is a functionof the deflection of the spring casing 200 and the drill bit 13.

The spring assembly 218 therefore can exert a restoring force on thedrill bit 13 in both the up-hole and down-hole directions. Thedual-action characteristic of the first and second springs 206, 208, itis believed, makes the spring assembly 218 more compact than acomparable spring assembly that employs multiple single-action springs.

Moreover, the spring assembly 18 is adapted for use under bothrelatively low and relatively high weight-on-bit conditions due to thecombined use of a relatively soft and a relatively hard spring. Inparticular, it is believed that Belleville washers of first (softer)spring 206 deflect (compress) when the weight-on-bit, i.e., thedown-hole force, on the drill bit 13 is relatively low. The Bellevillewashers of the second spring 208 do not deflect substantially under lowweight-on-bit conditions. The spring assembly 18 thus exerts arelatively low restoring force on the drill bit 13 under relatively lowweight-on-bit conditions. This feature permits axial vibrations of thedrill bit 13 to be transmitted to, and damped by the valve assembly 14.

Further increasing the weight-on-bit further compresses the Bellevillewashers of the first spring 206, until the Belleville washers of thefirst spring 206 become fully compressed. Additional increases in theweight-on-bit cause the Belleville washers of the second spring 208 todeflect (compress). The relatively high spring constant of the secondspring 208 increases the restoring force exerted by the spring assembly18 on the drill bit 13 as the Belleville washers of the second spring208 begin to deflect to deflect. The spring assembly 18 thus facilitatestransmission of axial vibrations to the valve assembly 16 under bothrelatively low and relatively high weight-on-bit conditions, whilepermitting axial vibration to be transmitted to and damped by the valveassembly 14 when the weight-on-bit is relatively low.

FIGS. 15A and 15B list equations for calculating the combined springconstant of the first and second springs 206, 208. Sample calculationscorresponding to a “soft” spring having a spring constant of 210K poundsper inch, and a “hard” spring having a spring constant of 1,160K poundsper inch also are presented. It should be noted that these particularvalues for the spring constants are provided for exemplary purposesonly, as the optimal spring constants for the first and second springs206, 208 are application-dependent.

The symbols listed in FIGS. 15A and 15B represent the followingparameters: k—spring constant; n—number of springs; h—free total heightof the spring; h_(w)—working height of the spring; k_(c)—total springconstant; L—total spring height;)L—maximum stroke; L_(e)—total springstack; *—spring deflection, where the subscripts “1” and “2” denote thesoft and hard springs, respectively.

The foregoing description is provided for the purpose of explanation andis not to be construed as limiting the invention. While the inventionhas been described with reference to preferred embodiments or preferredmethods, it is understood that the words which have been used herein arewords of description and illustration, rather than words of limitation.Furthermore, although the invention has been described herein withreference to particular structure, methods, and embodiments, theinvention is not intended to be limited to the particulars disclosedherein, as the invention extends to all structures, methods and usesthat are within the scope of the appended claims. Those skilled in therelevant art, having the benefit of the teachings of this specification,may effect numerous modifications to the invention as described herein,and changes may be made without departing from the scope and spirit ofthe invention as defined by the appended claims.

For example, FIGS. 16A and 16B depict a valve assembly 400. The valveassembly 400 is substantially identical to the valve assembly 16, withthe below-noted exceptions. (Components of the valve assembly 400 thatare substantially identical to those of the valve assembly 16 aredenoted herein by identical reference numerals.)

The valve assembly 400 comprises a valve casing 402. The valve casing402 has an inner surface 406. The inner surface 406 is tapered as shownin the figures. The taper of the inner surface 406 causes the innerdiameter of the valve casing 402 to decrease in the axial direction,between each end of the valve casing 402 and the approximate center ofthereof. In other words, the diameter of the valve casing 402 is maximalat the ends thereof, and is minimal at the approximate center thereof(in relation to the axial direction).

The valve assembly 400 also includes a coil mandrel 408 positionedwithin the housing, and movable in relation to the housing in the axial(“x”) direction. Outer surfaces 410 of the coil mandrel 408 are taperedin a manner similar those of the inner surface 406 of the casing 402.The outer surfaces 410 of the coil mandrel 408 and the inner surface 406of the casing 402 define a gap 412.

The taper of the inner surface 406 and the outer surface 410 causes thegap 412 to decrease in response to relative movement of the coil mandrel408 from the centered position depicted in FIG. 16A, the positiondepicted in FIG. 16B. Decreasing the gap 412 increases the resistance ofthe MRF to movement between the up-hole and down-hole chambers 150, 152.The damping force exerted by the valve assembly 400 therefore increaseswith the magnitude of the vibration of the drill bit 13. The decreasedgap also creates a higher electrical magnetic field within the MRF,thereby increasing the viscosity of the MRF.

FIGS. 17A and 17B depict another alternative embodiment of the valveassembly 16 in the form of a valve assembly 440. The valve assembly 440comprises a valve casing 442, and a mandrel 444 positioned within thevalve casing 442.

The mandrel 444 has an inner surface 446. A plurality of permanentmagnets 449 are embedded in the mandrel 444, proximate the inner surface446. (The valve assembly 440 does not include coils such as the coils136 of the valve assembly 14.)

The valve casing 442 includes a plurality inwardly-facing of projections450. Each projection is separated from the inner surface 446 of themandrel 444 by a gap 454 filled with MRF. The inner surface 446 isshaped so the gaps 454 are maximal when the mandrel 444 disposed in aneutral (centered) position in relation to the valve casing 442, asdepicted in FIG. 17A. The resistance offered by the MRF to relativemovement between the valve casing 442 and the coil mandrel 444 isminimal under this condition.

The inner surface 446 of the mandrel 444 is shaped so that axialmovement of the mandrel 444 from its neutral position decreases the gaps454, as shown in FIG. 17B. Moreover, the magnetic fields generated bythe permanent magnets 449 become focused in the gaps 454, therebyincreasing the flow resistance of the MRF in the gaps 450. Hence, theresistance of the MRF to relative movement between the coil mandrel 444and the valve casing 442 increases as the coil mandrel 444 moves fromits neutral position.

FIG. 18 depicts another alternative embodiment of the valve assembly 16in the form of a valve assembly 460. The valve assembly 460 comprises avalve casing 462, and a mandrel 464 positioned within the valve casing462.

The valve casing 462 and the mandrel 464 define a first, or up-holechamber 466 and a second, or down-hole chamber 468. The first and secondchambers 464, 466 are filled hydraulic fluid. The first and secondchambers 466, 468 are in fluid communication by way of a passage 470formed in the valve casing 462.

The valve assembly 460 also includes a valve 472 for restricting theflow of hydraulic fluid between the first and second chambers 464, 466by restricting a flow area of the passage 470. The valve 472 can becontrolled by a device such as the controller 146 to increase ordecrease the amount of restriction, and thus magnitude of the dampingforce produced by the valve assembly 460.

PARTS LIST

-   Vibration damping system 10-   Torsional bearing assembly 14-   Valve assembly 16-   Spring assembly 18-   Turbine-alternator module 20-   Outer casing 21 (turbine-alternator module 20)-   Drill pipe 22-   Bearing casing 50-   Bearing mandrel 52-   Radial bearing 54-   Ball bearings 55-   Centralizer feed-thru 56-   Locking pin 57-   Connector 58-   Connector 59-   Panel 60-   Passage 61-   Passage 63-   Inner surface 64 (of bearing mandrel 52)-   Grooves 70 (in bearing mandrel 52)-   Outer surface 72 (of bearing mandrel 64)-   Grooves 74 (in bearing casing 50)-   Inner surface 76 (of bearing casing 50)-   Passages 78-   Longitudinal centerline 80 (of grooves 70)-   Longitudinal centerline 82 (of bearing mandrel 52)-   LVDT 84-   Magnetic elements 86-   Sensor 88-   Compensation piston 90-   Up-hole side 90′ (of compensation piston 90)-   Down-hole side 90″ (of compensation piston 90)-   Reciprocating seals 91-   Grooves 92 formed compensation piston 90)-   Heal 93 (of seals 91)-   Lip 94-   Extension 95-   Groove 96 (in seals 91)-   Spring 97-   Valve casing 102 (of valve assembly 16)-   Outer casing 103 (of valve casing 102)-   Housing 104-   Coil mandrel 106-   First portion 108 (of housing 104)-   Second portion 110-   Third portion 112-   Radial bearing 120-   Sleeve 122-   Inner surface 124 (of coil mandrel 106)-   First linear bearing 125-   Second linear bearing 126-   Passage 127-   Recesses 128-   Outer surface portions 130-   Inner surface 132 (of second portion 110 of housing 104)-   Gap 135 (between inner surface 132 of second portion 110, and outer    surface portions 130 of coil mandrel 106)-   Coils 136-   Grooves 140 (in outer surface portions 130)-   Wireway 142-   Electrical feed thru 144-   Controller 146-   Alternator 147 (of turbine-alternator assembly)-   Magnetic fields (produced by coils 136)-   First (up-hole) chamber 150-   Second (down-hole) chamber 152-   Computing device 160 (of controller 146)-   Memory storage device 162-   Solid state relays 162-   Computer-executable instructions 164-   Spring casing 200 (of spring assembly 18)-   Spring mandrel 202-   Spring stack 205-   First spring 206-   Second spring 208-   Inner surface 209 (of spring mandrel 202)-   Passage 210-   Coupling 211-   First spacer 212-   Second spacer 214-   Lip 216 (of spring casing 200)-   Third spacer 218-   Compensation module 300-   Casing 302 (of compensation module 300)-   Mandrel 304-   Radial bearing 305-   Sliding compensation piston 306-   Passage 314 (in mandrel 304)-   Inner surface 315-   Chamber 316-   Valve assembly 400-   Valve casing 402-   Inner surface 406 (of valve casing 402)-   Coil mandrel 408-   Outer surfaces 410 (of coil mandrel 408)-   Gap 412-   Valve assembly 440-   Valve casing 442-   Coil mandrel 444-   Inner surface 446 (of coil mandrel 444)-   Magnets 449-   Projections 450 (on valve casing 442)-   Gap 454 (between inner surface 446 and projections 450)-   Valve assembly 460-   Valve casing 462-   Mandrel 464-   Casing 462-   First chamber 466-   Second chamber 468-   Passage 470 (between first and second chambers 466, 468)-   Valve 472

1-38. (canceled)
 39. A vibration damping system for use in a drillstring for drilling a drill hole, comprising: a bearing comprising afirst member and a second member coupled to the first member so that thefirst member can translate in an up-hole and a down-hole direction inrelation to the second member and torque can be transferred between thefirst and the second members; a valve assembly comprising a first membersecurely coupled to the first member of the torsional bearing assemblyso that the first member of the valve assembly translates in the up-holeand down-hole directions with the first member of the torsional bearingassembly; a second member securely coupled to the second member of thetorsional bearing assembly so that the second member of the valveassembly translates in the up-hole and down-hole directions with thesecond member of the first torsional bearing assembly, the first andsecond members of the valve assembly defining a first and a secondchamber for holding a supply of a fluid so that the fluid flows betweenthe first and the second chambers in response to relative movementbetween the first and second members of the valve assembly; and meansfor varying a flow resistance of the fluid; and a spring assemblycomprising a first member securely coupled to the first member of thevalve assembly so that the first member of the spring assemblytranslates in the up-hole and down-hole directions with the first memberof the valve assembly; a second member securely coupled to the secondmember of the valve assembly the so that the second member of the springassembly translates in the up-hole and down-hole directions with thesecond member of the valve assembly; and a spring for resisting relativemovement between the first and second members of the spring assembly.40-43. (canceled)